The present invention relates to an automatic air extraction hydraulic diaphragm pump driven by a cam mechanism, and more particularly to a pulsation free pump provided with a pulsation adjustable feature capable of adjusting variable pulsation of a pump discharge liquid to be minimum during operations of the pumps.
Twin pulsation free pumps are illustrated in FIGS. 1A and 1B and have generally been known. In FIG. 1A, two automatic air extraction hydraulic diaphragm pumps P1 and P2 are arranged in parallel to each other to be driven by a cam mechanism 10 at a phase difference of 180.degree.. In FIG. 1B, two automatic air extraction hydraulic diaphragm pumps P1 and P2 are arranged in series to be driven by a cam mechanism 10 at a phase difference of 180.degree.. The first and second pumps P1 and P2 have pump chambers 12, 12 which are coupled to a single suction inlet pipe 14 and to a discharge pipe 16. The first and second pumps P1 and P2 also have hydraulic chambers 18, 18 which are coupled through automatic air extractors 20, 20 to oil reservoirs 22, 22.
Similarly, triple pulsation free pumps may be constituted by three automatic air extraction hydraulic diaphragm pumps driven by a cam mechanism at a phase difference of 120.degree..
FIG. 2A is illustrative of a property in a theoretical discharge flow rate of the above twin pulsation free pumps. FIG. 2B is illustrative of a property in an actual composite discharge flow rate of the above twin pulsation free pumps. FIG. 3A is illustrative of a property in a theoretical discharge flow rate of the above triple pulsation free pumps. FIG. 3B is illustrative of a property in an actual composite a discharge flow rate of the above triple pulsation free pumps.
The conventional pulsation free pumps described above, however, exhibit undesirable pulsation discharge flows due to the following five factors.
The first factor is concerned with a clearance in the driving section. The second factor is concerned with a residual air in a hydraulic driving section. The third factor is concerned with a leakage of the liquid in an air extraction process. The fourth factor is concerned with a residual air in a pump operation section. The fifth factor is concerned with a leakage of the liquid from a check valve. Due to the above factors, the first pump P1 positioned to follow the second pump P2 shows a delay (.DELTA.t) in time of the discharge as well as a loss (.DELTA.q) of the discharge flow rate. The following descriptions will focus on each of the influences caused by the above five factors.
As to the first factor concerned with the clearance in the driving section, even if a clearance exists in a rotation driving section, no change in the discharge flow rate appears due to the clearance being unidirectional. If, however, a clearance exists in a reciprocal driving section then the direction of the clearance is different between the discharge and suction processes thereby a waveform of the actual discharge flow rate of the first pump P1 is shifted from the theoretical waveform thereof in a direction of delay as illustrated in FIG. 4A. Particularly, the clearance direction is changed at a time .theta..sub.3 when the first pump P1 enters into a suction process. As a result, the composite discharge flow rate is reduced at a time when the first pump P1 initiates the discharge as well as increased at a time when the first pump P1 initiates the suction and the second pump P2 initiates the discharge as illustrated in FIG. 4B.
As to the second factor concerned with the influences of the residual air in the hydraulic driving section, at a time .theta..sub.0 when the first pump P1 enters into the suction process, an air pressure is raised thereby causing an undesirable and additional time consumption for obtaining the required discharge pressure. An increase of the discharge flow rate of the first pump P1 has a time delay (.DELTA.t1) as illustrated in FIG. 5A. The composite discharge flow rate has a certain loss (.DELTA.q1) of the discharge flow rate as illustrated in FIG. 5B.
As to the third factor concerned with the leakage of the liquid in the air extraction process, at the time .theta..sub.0 when the first pump P1 enters into the discharge process, a small amount of the oil liquid is unwillingly extracted from the hydraulic driving section during air extraction. Such oil leakage leads to an undesirable and additional time consumption for obtaining the required discharge pressure of the first pump P1 thereby an increase of the discharge flow rate has a time delay (.DELTA.t2) as illustrated in FIG. 6A. As a result, the composite discharge flow rate has a certain loss (.DELTA.q1) of the discharge at the time when the first pump P1 initiates the discharge as illustrated in FIG. 6B.
As to the fourth factor concerned with the influences by the residual air in the pump operation section, at the time .theta..sub.0 when the first pump P1 enters into the discharge process, an air pressure is raised thereby causing an undesirable and additional time consumption for obtaining the required discharge pressure. An increase of the discharge flow rate of the first pump P1 has a time delay (t1) as illustrated in FIG. 5A. The composite discharge flow rate has a certain loss (.DELTA.q1) of the discharge flow rate as illustrated in FIG. 5B.
As to the fifth factor concerned with the influences due to the leakage of the liquid from the check valve, when a leakage of the liquid is generated from the check valve positioned at the discharge side of the first pump P1, then during the discharge process of the first pump P1 there appears a leakage of the discharge liquid from the inside of the first pump P1 into the suction inlet pipe thereby the discharge flow rate of the first pump P1 is totally reduced as illustrated in FIG. 7A. As a result, the composite discharge flow rate is reduced from the theoretical discharge flow rate during the discharge process of the first pump P1 as illustrated in FIG. 7B.
When a leakage of the liquid is generated at the check valve positioned at the suction side of the first pump P1, then during the discharge process of the first pump P1, the discharge liquid flows during the discharge process of the first pump P1 in a reverse direction from the discharge pipe into the inside of the first pump P1 thereby a suction flow rate of the first pump p1 is totally reduced as illustrated in FIG. 7C. As a result, the composite discharge flow rate is reduced from the theoretical discharge flow rate during the suction process of the first pump P1 namely during the discharge process of the second pump P2 as illustrated in FIG. 7D.
The above problems caused by the first and fifth factors may readily be settled by a certain design change of pump elements, whereas settlements of the problems caused by the remaining factors, namely, the second, third and fourth factors would be difficult. There has been proposed a compensation of cams in the cam mechanism 10 for settlements of the above problems due to the above second, third and fourth factors. The compensations already proposed may be classified into three types as follows.
The first proposal is to compensate the cams for changing the discharge property in an initiation stage of the discharge process. The cams in the cam mechanism 10 illustrated in FIGS. 1A and 1B are compensated in those shapes so that the discharge flow rate property is set at a waveform represented by the real line in FIG. 8B. Whereas the pulsation is generated in the discharge initiation stage, a removal of the pulsation from the composite discharge flow rate in the discharge actually follows the completion in compression of the residual air represented by crosshatching in FIG. 8B. The pulsation of the composite discharge flow rate could not be removed.
The second proposal is to make a cam compensation for placing the cam in discharge allowable state prior to the actual discharge timing. The shape of the cam is compensated so that the pump discharge flow rate has a waveform represented by a real line in FIG. 9B. A compression corresponding to a volume, represented by a crosshatched portion in FIG. 9B, associated with the discharge flow rate is set prior to the actual discharge, for which reason on the discharge initiation stage the discharge have already been suitable due to including an extra discharge flow rate. Such extra discharge flow rate may, however, cause an increase in the composite discharge flow rate thereby the pulsation is generated as illustrated in FIG. 9.
The third proposal is to place the cam in discharge allowable state prior to the actual discharge timing so as to set the discharge flow rate at zero in the actual discharge initiation. The shape of the cams is compensated so that the discharge flow rate has a waveform represented by a real line in FIG. 10. A compression corresponding to a volume, represented by a crosshatched portion, of the pump discharge flow rate is set prior to the actual discharge so that the discharge has already been suitable on the discharge initiation stage. At the initiation of the discharge, the discharge flow rate is set at zero so that no extra discharge flow rate is generated thereby the composite discharge flow rate has no pulsation as illustrated in FIG. 10A. Namely, the shape of the cams may be compensated so that the pulsation is removed in the discharge initiation stage.
The influences due to the above second, third and fourth factors may be resolved by making the compensation in shape of the cams to obtain pulsation free flow. In other words, the pulsation may be removed only by compensation in the shape of the cams.
As described above, the convention pulsation free pump shows the reduction in the composite discharge flow rate in its initial stage due to the residual air in the hydraulic driving section, the liquid leakage in the air extraction process and the residual air in the pump operation section and others, for which reason the reduction would be unavoidable. The unavoidable reduction may be compensated by the extra discharge flow rate to ensure the desirable pulsation free discharge flow rate.
A magnitude of the reduction of the discharge flow rate depends upon the operation conduction of the pump system such as a discharge pressure and pipe lines, while a magnitude of the extra discharge flow rate is free from such condition. To ensure the pulsation free discharge flow rate, it is necessary to adjust an amount of the extra discharge flow rate for compensations for such variable reductions of the discharge flow rate due to the variable pump conditions.
In the conventional pulsation free pumps, the adjustment for ensuring the pulsation free discharge flow rate would substantially be impossible on the ground that the compensation in the shape of the cams would be restricted and a variation in an angular velocity is limited as being substantially defined by a stepping motor. Namely, the conventional cam shape compensation method is insufficient to exactly remove the pulsation from the discharge flow rate in response to largely variable pump operation conditions.
The conventional automatic air extraction hydraulic diaphragm pumps have a structure as illustrated in FIG. 11A. Inside a diaphragm pump body 40, there is provided a hydraulic chamber 44 and a pump chamber 46 which are separated by a diaphragm 42. The hydraulic chamber 44 is provided with plunger 48 penetrating the hydraulic chamber 44. The pump chamber 46 is provided with a suction port 54 and a discharge port 56 via check valves 50 and 52 respectively. A reciprocating operation of the plunger 48 causes a variation in pressure of the oil in the hydraulic chamber 44 thereby the diaphragm shows a pulse oscillation motion which allows the pump chamber 46 to show the pump operation.
On a top portion of the diaphragm pump body 40, an oil reservoir 58 is provided wherein an oil reserving chamber 60 within the oil reservoir 58 is connected to the above hydraulic chamber 44 through a multi-function valve 60a and an oil passage 64 provided in the oil reservoir 58 as well as through an oil passage 66 provided in the diaphragm pump body 40. As a result, the above multi-function valve 60a is so operated as to supplement the oil to the hydraulic chamber 44 when the hydraulic chamber 44 is deficient in oil due to operations of the plunger 48 and further to have the oil discharge from the hydraulic chamber 44 into the oil reservoir when the hydraulic chamber 44 has an excess of the oil. The above multi-function valve 60a is capable both of an air extraction for discharge of bubbles generated in the hydraulic chamber 44 by operations of the plunger 48 and of a supplement of a driving oil for compensation for a reduction thereof due to a leakage from the hydraulic chamber 44. Further, there is provided a safety valve 60b for allowing, in the hydraulic chamber 44, an escape of the excess of the oil pressure over a regulation valve.
There is further provided a piston pump 72 for driving a piston 70 via a cam 68 showing a rotation driving which is synchronized with a reciprocal motion of the plunger 48. The multi-function valve 60a is coupled to the piston pump 72 via the pump chamber 74 and an oil feeder pipe 76 to force the multi-function valve 60a to show opening and closing operations in association with the pump operation of the piston pump 72.
As illustrated in FIG. 11B, the above multi-function valve 60a has a valve body 80 within which there is formed a pressure chamber 82 which is coupled via an oil passage 84 to the oil feeder pipe 76 extending from the piston pump 72. On a top of the valve body 80, a flow rate adjuster 86 comprising an orifice is provided to introduce the oil in the oil reservoir 60 into the pressure chamber 82. A piston 88 is inserted into and supported by the pressure chamber 82 so that the piston 88 is fixed at an intermediate position of the pressure chamber 82. On the bottom of the valve body 80, there is formed a stem 90 into which inserted is a valve stem 94 which is closed by a spring 92. The valve stem 94 extends to penetrate the pressure chamber 82 and a portion thereof projecting from the pressure chamber is united with a valve section 98 having a tapered shape 96. The above multi-function valve 60a allows the oil to be fed discontinuously by the piston pump 72 to thereby generate a pressure difference when the pressured oil passes the orifice of the flow rate adjuster 86. The pressure difference may cause the piston 88 pressed down to have the stem 90 open for oil supplement into the hydraulic chamber 44 of the diaphragm pump body 40 together with an extraction of the air generated in the hydraulic chamber 44.
The above multi-function valve 60a may comprise a differential pressure automatic air extraction ball valve as illustrated in FIG. 12. In FIG. 12, the differential pressure automatic air extraction ball valve 30 is provided at its top portion with an adjusting nut 31, a valve body 34 with a seat 33 for a ball 32, and a bottom screw 35 engaged with the above adjusting nut 31 wherein the screw is inserted into the valve body 34. The valve 30 is further provided with an adjustable pipe 36 with a top seat 37 for the ball 32, and a stopper nut 38 engaged with the screw 35 of the above adjustable pipe 36.
The above differential pressure automatic air extraction ball valve 30 is so constructed that the ball 32 moves from top to bottom in the initiation of the pump suction process thereby a small amount of the oil flows from the oil reservoir 60 into the hydraulic chamber 44 and further the ball 32 moves from bottom to top in an initiation of the pump discharge process thereby a small amount of the oil together with an air in the hydraulic chamber 44 is discharged flow rate of the pressured oil is set larger than the suction flow rate since a pressure difference of the oils in discharge between inside the pump chamber 46 and an atmosphere is larger than that of the oil in suction.